Two-stage pressurising refrigeration cycle device

ABSTRACT

A refrigerant discharge capacity of a high-pressure side compression mechanism and a refrigerant discharge capacity of a low-pressure side compression mechanism can be independently controlled, in a two-stage pressurizing refrigeration cycle device. The refrigerant discharge capacity of the low-pressure side compression mechanism is determined based on an outside air temperature, an air temperature at the evaporator, and a preset temperature. Furthermore, the refrigerant discharge capacity of the high-pressure side compression mechanism is determined based on the determined refrigerant discharge capacity of the low-pressure side compression mechanism such that an effective capacity ratio is not less than 1 nor more than 3. Therefore, in the two-stage pressurizing refrigeration cycle device, COP can be improved with a simple structure and control.

CROSS REFERENCE TO RELATED APPLICATION

This application is based on Japanese Patent Application No. 2010-154680 filed on Jul. 7, 2010, the contents of which are incorporated herein by reference in its entirety.

TECHNICAL FIELD

The present invention relates to a two-stage pressurizing refrigeration cycle device including a low-pressure side compression mechanism and a high-pressure side compression mechanism for increasing the pressure of refrigerant through multiple stages.

BACKGROUND ART

Conventionally, Patent Document 1 discloses a two-stage pressurizing refrigeration cycle device for increasing the pressure of refrigerant in multiple stages. The refrigeration cycle device includes a low-pressure side compression mechanism for compressing and discharging the low-pressure refrigerant into an intermediate-pressure refrigerant, and a high-pressure side compression mechanism for compressing and discharging the intermediate-pressure refrigerant discharged from the low-pressure side compression mechanism, into a high-pressure refrigerant therefrom.

More specifically, the two-stage pressurizing refrigeration cycle device disclosed in Patent Document 1 includes a radiator for dissipating heat from the high-pressure refrigerant discharged from the high-pressure side compression mechanism, and an intermediate-pressure expansion valve for decompressing and expanding a part of the high-pressure refrigerant flowing from the radiator into the intermediate-pressure refrigerant. The two-stage pressurizing refrigeration cycle device is the so-called economizer refrigeration cycle device that guides the intermediate-pressure refrigerant decompressed by the intermediate-pressure expansion valve to a suction side of the high-pressure side compression mechanism.

In this kind of economizer refrigeration cycle device, the high-pressure side compression mechanism can suck a mixture of the intermediate-pressure refrigerant decompressed by the intermediate-pressure expansion valve, and the intermediate-pressure refrigerant discharged from the low-pressure side compression mechanism. Thus, the low-temperature mixed refrigerant can be sucked into the high-pressure side compression mechanism, and thereby it can improve the compression efficiency of the high-pressure side compression mechanism, as compared to the case in which only the intermediate-pressure refrigerant discharged from the low-pressure side compression mechanism is sucked.

Further, in the economizer refrigeration cycle device, the high-pressure side compression mechanism and the low-pressure side compression mechanism have substantially the same compression ratio to each other. The pressure of the intermediate-pressure refrigerant (intermediate refrigerant pressure) is designed to approach a target intermediate refrigerant pressure defined as a geometric mean between the pressure of the high-pressure refrigerant (high-pressure side refrigerant pressure) and the pressure of the low-pressure refrigerant (low-pressure side refrigerant pressure), thereby improving the coefficient of performance (COP) of the cycle.

In the two-stage pressurizing refrigeration cycle device disclosed in Patent Document 1, a throttle opening degree of the intermediate-pressure expansion valve is changed such that the intermediate refrigerant pressure approaches the target intermediate refrigerant pressure to thereby improve the COP.

RELATED ART DOCUMENTS Patent Documents [Patent Document 1]

-   Japanese Unexamined Patent Publication No. 2006-242557

The two-stage pressurizing refrigeration cycle device disclosed in Patent Document 1, however, determines the target intermediate refrigerant pressure based on both the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure. Thus, pressure detection means must be provided for detecting both the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure, and thereby the manufacturing costs of the two-stage pressurizing refrigeration cycle device may be increased.

When the throttle opening degree of the intermediate-pressure expansion valve is changed so as to cause the intermediate refrigerant pressure to approach the target intermediate refrigerant pressure, the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure may be changed, thereby complicating the control for stabilizing (converging) a refrigerant discharge capacity of each compression mechanism.

The intermediate-pressure refrigerant flowing from the intermediate-pressure expansion valve would be often converted into a liquid-phase state or a gas-liquid two-phase state only by changing the throttle opening of the intermediate-pressure expansion valve such that the intermediate refrigerant pressure approaches the target intermediate refrigerant pressure. As a result, there poses the problem of liquid compression that the high-pressure side compression mechanism forcedly compresses an incompressible fluid, thereby degrading the reliability of the high-pressure side compression mechanism, that is, the reliability of the entire two-stage pressurizing refrigeration cycle device.

BRIEF SUMMARY OF THE INVENTION

The present invention has been made in view of the above points, and it is a first object of the present invention to provide a two-stage pressurizing refrigeration cycle device which can improve the COP with a simple structure and control.

It is a second object of the present invention to provide a two-stage pressurizing refrigeration cycle device that can provide high reliability with a simple structure and control.

To achieve the object, a two-stage pressurizing refrigeration cycle device according to a first aspect of the present invention includes: a low-pressure side compression mechanism, which compresses a low-pressure refrigerant into an intermediate-pressure refrigerant, and discharges the compressed refrigerant therefrom; a high-pressure side compression mechanism, which compresses the intermediate-pressure refrigerant discharged from the low-pressure side compression mechanism into a high-pressure refrigerant to discharge the compressed refrigerant therefrom; a radiator, which exchanges heat between outdoor air and the high-pressure refrigerant discharged from the high-pressure side compression mechanism, to dissipate heat from the refrigerant; an intermediate-pressure expansion valve, which decompresses and expands the high-pressure refrigerant flowing from the radiator into an intermediate-pressure refrigerant to flow the intermediate-pressure refrigerant into a suction side of the high-pressure side compression mechanism; a low-pressure expansion valve, which decompresses and expands the high-pressure refrigerant flowing from the radiator into a low-pressure refrigerant; and an evaporator, which evaporates the low-pressure refrigerant decompressed and expanded by the low-pressure expansion valve by exchanging heat with air to be blown into a space for cooling, and causes the evaporated refrigerant to flow into a suction side of the low-pressure side compression mechanism. Furthermore, the two-stage pressurizing refrigeration cycle device includes a first discharge capacity controller and a second discharge capacity controller. The first discharge capacity controller is configured to determine a refrigerant discharge capacity of at least one of the high-pressure side compression mechanism and the low-pressure side compression mechanism, such that the refrigerant discharge capacity is increased in accordance with an increase in at least one of an outside air temperature of the outdoor air for exchanging heat with the high-pressure refrigerant at the radiator and an air temperature of air for exchanging heat with the low-pressure refrigerant at the evaporator. The second discharge capacity controller is configured to determine a refrigerant discharge capacity of the other compression mechanism of the high-pressure side compression mechanism and the low-pressure side compression mechanism, based on the determined refrigerant discharge capacity of the one compression mechanism. In addition, the second discharge capacity controller determines the refrigerant discharge capacity of the other compression mechanism, such that an effective capacity ratio defined as N2×V2/N1×V1 is within a predetermined reference range, when V1 is a discharge capacity of the high-pressure side compression mechanism, N1 is the number of revolutions of the high-pressure side compression mechanism, V2 is a discharge capacity of the low-pressure side compression mechanism, and N2 is the number of revolutions of the low-pressure side compression mechanism.

Because the first discharge capacity controller determines the refrigerant discharge capacity of the one compression mechanism based on at least one of the outside air temperature and the air temperature at the evaporator, and the second discharge capacity controller determines the refrigerant discharge capacity of the other compression mechanism based on the determined refrigerant discharge capacity of the one compression mechanism, the refrigerant discharge capacities of the respective compression mechanisms can be readily determined.

At this time, the second discharge capacity controller determines the refrigerant discharge capacity of the other compression mechanism such that the effective capacity ratio is within a predetermined reference range. Thus, only by appropriately setting the reference range, the intermediate refrigerant pressure can approach the value substantially corresponding to the geometric mean between the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure.

Thus, the COP of the two-stage pressurizing refrigeration cycle device can be improved with a simple structure without the need for expensive pressure detection means under extremely easy control.

Regardless of the refrigerant discharge capacity of each compression mechanism, the throttle opening degree of the intermediate-pressure expansion valve can be determined to convert the intermediate-pressure refrigerant flowing from the intermediate-pressure expansion valve into a gas phase. Thus, the problem of liquid compression at the high-pressure side compression mechanism can be avoided.

Thus, the invention can improve the reliability of the high-pressure side compression mechanism with a simple structure, that is, the reliability of the entire two-stage pressurizing refrigeration cycle device. The term “discharge capacity of the compression mechanism” as used herein means a theoretical discharge flow rate per rotation of the compression mechanism, specifically, a flow rate calculated geometrically.

For example, in a two-stage pressurizing refrigeration cycle device according to a second aspect of the present invention, the intermediate-pressure expansion valve decompresses and expands one high-pressure refrigerant flow branched at a branching portion in which the high-pressure refrigerant flowing from the radiator is branched, and the low-pressure expansion valve decompresses and expands the other high-pressure refrigerant flow branched at the branching portion. The refrigeration cycle device further includes an intermediate heat exchanger in which heat is exchanged between the low-pressure refrigerant decompressed and expanded by the intermediate-pressure expansion valve and the other high-pressure refrigerant branched by the branching portion.

The two-stage pressurizing refrigeration cycle device of the invention includes an intermediate heat exchanger, which can heat the intermediate-pressure refrigerant flowing from the intermediate-pressure expansion valve to easily convert the refrigerant into the gas-phase refrigerant. As a result, the reliability of the two-stage pressurizing refrigeration cycle device can be surely improved.

Because the other high-pressure refrigerant flow branched at a branching portion is cooled to thereby enlarge a difference in enthalpy between the refrigerants at inlet and outlet of the evaporator, a refrigeration capacity exhibited by the evaporator can be increased. As a result, the COP of the two-stage pressurizing refrigeration cycle device can be more improved.

For example, in a two-stage pressurizing refrigeration cycle device according to a third aspect of the present invention, the one compression mechanism is the low-pressure side compression mechanism, and the other compression mechanism is the high-pressure side compression mechanism.

Thus, the first discharge capacity controller can determine the refrigerant discharge capacity of the low-pressure side compression mechanism based on at least one of the outside air temperature and the air temperature at the evaporator to thereby directly control a refrigerant evaporation pressure of the evaporator. Therefore, the air temperature blown into a space for cooling tends to be adjusted to a desired temperature.

In the two-stage pressurizing refrigeration cycle device according to a fourth aspect of the present invention, when an absolute value of a difference in temperature between the air temperature at the evaporator and a target cooling temperature of the space for cooling is equal to or less than a predetermined reference temperature difference, the second discharge capacity controller determines the refrigerant discharge capacity of the other compression mechanism based on the refrigerant discharge capacity of the one compression mechanism.

In this case, the second discharge capacity controller can control the refrigerant discharge capacity of the other compression mechanism, regardless of the refrigerant discharge capacity of the one compression mechanism, until the absolute value of a difference between the air temperature at the evaporator and the target cooling temperature of the space for cooling is equal to or less than a predetermined reference temperature difference. Thus, for example, an operation mode for rapidly cooling the space for cooling can be executed upon the start-up of the two-stage pressurizing refrigeration cycle device.

In the two-stage pressurizing refrigeration cycle device according to a fifth aspect of the present invention, when the air temperature at the evaporator is higher than the target cooling temperature, and when the absolute value of the difference in temperature between the air temperature at the evaporator and the target cooling temperature of the space for cooling is equal to or less than the predetermined reference temperature difference, the second discharge capacity controller may determine the refrigerant discharge capacity of the other compression mechanism based on the refrigerant discharge capacity of the one compression mechanism. In this case, the capacity control of both the compression mechanisms can be effectively performed.

In the two-stage pressurizing refrigeration cycle device according to a sixth aspect of the present invention, each of the high-pressure side compression mechanism and the low-pressure side compression mechanism may be configured by a fixed displacement compression mechanism having a fixed discharge capacity. Furthermore, refrigeration cycle device may further include a high-pressure side electric motor which rotatably drives the high-pressure side compression mechanism, and a low-pressure side electric motor which rotatably drives the low-pressure side compression mechanism. In addition, the number of revolutions of the high-pressure side electric motor and the number of revolutions of the low-pressure side electric motor may be independently controllable.

In this case, the discharge capacity of the high-pressure side compression mechanism and the discharge capacity of the low-pressure side compression mechanism respectively become constant, so that the effective capacity ratio can be easily set within the reference range only by changing at least one of the number of revolutions of the high-pressure side compression mechanism and the number of revolutions of the low-pressure side compression mechanism.

Furthermore, in the two-stage pressurizing refrigeration cycle device according to a seventh aspect of the present invention, each of the high-pressure side compression mechanism and the low-pressure side compression mechanism may be configured by a variable displacement compression mechanism having a variable discharge capacity, and the discharge capacity of the high-pressure side compression mechanism and the discharge capacity of the low-pressure side compression mechanism may be independently controllable.

In this case, the discharge capacity of the high-pressure side compression mechanism and the discharge capacity of the low-pressure side compression mechanism can be independently changed. Even if the numbers of revolutions of both the compression mechanisms are set to the same level, the effective capacity ratio can be easily set within the reference range. Thus, both the compression mechanisms can be driven by common driving means.

Furthermore, in the above any two-stage pressurizing refrigeration cycle device, the second discharge capacity controller may determine the refrigerant discharge capacity of the other compression mechanism, such that the effective capacity ratio satisfies the following formula:

1≦N2×V2/N1×V1≦3.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is an entire configuration diagram of a two-stage pressurizing refrigeration cycle device in a first embodiment;

FIG. 2 is a flowchart showing a control process of the two-stage pressurizing refrigeration cycle device in the first embodiment;

FIG. 3( a) is a graph showing a change in COP ratio against a change in effective capacity ratio under a predetermined condition A, and FIG. 3( b) is a graph showing a change in COP ratio against a change in effective capacity ratio under another condition B; and

FIG. 4 is an entire configuration diagram of a two-stage pressurizing refrigeration cycle in a second embodiment.

EMBODIMENTS FOR CARRYING OUT THE INVENTION First Embodiment

Referring to FIGS. 1 to 3, a first embodiment of the invention will be described below. FIG. 1 shows an entire configuration diagram of a two-stage pressurizing refrigeration cycle device 10 of this embodiment. The two-stage pressurizing refrigeration cycle device 10 is applied to a refrigeration machine, and serves to cool the blast air blown into a freezer as a space for cooling down to an ultralow temperature of, for example, about −30 to −10° C.

As shown in FIG. 1, the two-stage pressurizing refrigeration cycle device 10 includes two compressors, namely, a high-pressure side compressor 11 and a low-pressure side compressor 12. The cycle device 10 serves to increase the pressure of the refrigerant circulating through the cycle in multiple stages. The refrigerant used can be normal fluorocarbon refrigerant (for example, R404A). Refrigeration machine oil (oil) for lubricating sliding portions of the low-pressure side compressor 12 and the high-pressure side compressor 11 is mixed into the refrigerant, so that a part of the refrigeration machine oil circulates through the cycle together with the refrigerant.

The low-pressure side compressor 12 is an electric compressor including a low-pressure side compression mechanism 12 a for compressing the low-pressure refrigerant into the intermediate-pressure refrigerant and discharging the compressed intermediate-pressure refrigerant therefrom, and a low-pressure side electric motor 12 b for rotatably driving the low-pressure side compression mechanism 12 a. The low-pressure side compression mechanism 12 a is comprised of a fixed displacement compression mechanism whose discharge capacity V2 is fixed. Specifically, the low-pressure side compression mechanism 12 a can employ various types of compression mechanisms, including a scroll compression mechanism, a vane compression mechanism, a rolling piston compression mechanism, and the like.

The low-pressure side electric motor 12 b is an AC motor whose operation (number of revolutions) is controlled by an alternating current output from a low-pressure side inverter 22. The low-pressure side inverter 22 outputs the AC with a frequency corresponding to a control signal output from a refrigeration machine controller 20 to be described later. Under control of the frequency, the refrigerant discharge capacity of the low-pressure side compressor 12 (specifically, the low-pressure side compression mechanism 12 a) is changed.

Thus, in this embodiment, the low-pressure side electric motor 12 b serves as discharge capacity changing means for the low-pressure side compressor 12. Apparently, the low-pressure side electric motor 12 b may employ a DC motor whose number of revolutions is controlled by a control voltage output from the refrigeration machine controller 20. A discharge port of the low-pressure side compressor 12 (specifically, the low-pressure side compression mechanism 12 a) is coupled to a suction port of the high-pressure side compressor 11.

The high-pressure side compressor 11 has the same basic structure as that of the low-pressure side compressor 12. Thus, the high-pressure side compressor 11 is an electric compressor comprised of a high-pressure side compression mechanism 11 a and a high-pressure side electric motor 11 b. The high-pressure side compression mechanism 11 a is adapted for compressing the intermediate-pressure refrigerant discharged from the low-pressure side compressor 12 into a high-pressure refrigerant and discharging the compressed high-pressure refrigerant therefrom.

The high-pressure side compression mechanism 11 a is comprised of a fixed displacement compression mechanism whose discharge capacity V1 is fixed. The high-pressure side electric motor 11 b has its number of revolutions controlled by the alternating current output from a high-pressure side inverter 21. The compression ratio of the high-pressure side compression mechanism 11 a is substantially the same as that of the low-pressure side compression mechanism 12 a in this embodiment.

A discharge port of the high-pressure side compressor 11 (specifically, the high-pressure side compression mechanism 11 a) is coupled to a refrigerant inlet side of a radiator 13. The radiator 13 is a heat exchanger for heat dissipation that exchanges heat between the high-pressure refrigerant discharged from the high-pressure side compressor 11 and an air outside the freezer (outdoor air) blown by a cooling fan 13 a to dissipate heat from the high-pressure refrigerant so as to cool the refrigerant.

The cooling fan 13 a is an electric blower whose number of revolutions (the amount of blast air) is controlled by a control voltage output from the refrigeration machine controller 20. The two-stage pressurizing refrigeration cycle device 10 of this embodiment forms a subcritical refrigeration cycle in which the high-pressure side refrigerant pressure does not exceed a critical pressure of the refrigerant using a fluorocarbon refrigerant as the refrigerant. Thus, the radiator 13 serves as a condenser for condensing the refrigerant.

A branching portion 14 for branching the flow of refrigerant flowing thereinto from the radiator 13 is coupled to a refrigerant outlet of the radiator 13. The branching portion 14 has a three-way joint structure with three inlet/outlet ports. One of the inlet/outlet ports serves as a refrigerant inlet, and two of them serve as a refrigerant outlet. Such a branching portion 14 may be formed by connecting pipes, or may be formed by providing a plurality of refrigerant passages in a metal or resin block.

One of the refrigerant outlets of the branching portion 14 is coupled to the inlet side of an intermediate-pressure expansion valve 15, and the other outlet of the branching portion 14 is coupled to the inlet side of a high-pressure refrigerant flow path 16 a of an intermediate heat exchanger 16. The intermediate-pressure expansion valve 15 is a thermal expansion valve that compresses and expands the high-pressure refrigerant flowing from the radiator 13 into the intermediate-pressure refrigerant.

More specifically, the intermediate-pressure expansion valve 15 has a temperature sensing portion disposed on the outlet side of an intermediate-pressure refrigerant flow path 16 b of the intermediate heat exchanger 16. The expansion valve 15 senses a degree of superheat of the refrigerant on the outlet side of the intermediate-pressure refrigerant flow path 16 b based on the temperature and pressure of the refrigerant on the outlet side of the intermediate-pressure refrigerant flow path 16 b. The expansion valve 15 adjusts the opening degree of the valve (refrigerant flow rate) by a mechanical mechanism such that the superheat degree becomes a predetermined value preset. The outlet side of the intermediate-pressure expansion valve 15 is coupled to the inlet side of the intermediate-pressure refrigerant flow path 16 b.

The intermediate heat exchanger 16 exchanges heat between an intermediate-pressure refrigerant flowing through the intermediate-pressure refrigerant flow path 16 b and decompressed and expanded by the intermediate-pressure expansion valve 15, and the other high-pressure refrigerant flowing through the high-pressure refrigerant flow path 16 a and branched by the branching portion 14. The high-pressure refrigerant is decompressed to have its temperature decreased. Thus, the intermediate heat exchanger 16 heats the intermediate-pressure refrigerant flowing through the intermediate-pressure refrigerant flow path 16 b, and cools the high-pressure refrigerant flowing through the high-pressure refrigerant flow path 16 a.

Specifically, the intermediate heat exchanger 16 employs a plate-type heat exchanger including a plurality of heat-transfer plates laminated, and the intermediate-pressure refrigerant flow paths 16 b and the high-pressure refrigerant flow paths 16 a which are alternatively arranged between the respective heat-transfer plates. The heat exchanger is designed to exchange heat between the high-pressure refrigerant and the intermediate-pressure refrigerant via the heat-transfer plates.

Alternatively, the heat exchanger 16 may employ a double-pipe heat exchanger structure comprised of an outside pipe forming the high-pressure refrigerant flow path 16 a, and an inside pipe forming the intermediate-pressure refrigerant flow path 16 b located inside the flow path 16 a. Apparently, the high-pressure refrigerant flow path 16 a may be formed as the inside pipe, and the intermediate-pressure refrigerant flow path 16 b as the outside pipe. Further, refrigerant pipes forming the high-pressure refrigerant flow path 16 a and the intermediate-pressure refrigerant flow path 16 b may be coupled to each other to exchange heat therebetween.

The intermediate heat exchanger 16 shown in FIG. 1 employs a parallel heat exchanger in which the flow direction of the high-pressure refrigerant flowing through the high-pressure refrigerant flow path 16 a is the same as that of the intermediate-pressure refrigerant flowing through the intermediate-pressure refrigerant flow path 16 b. Alternatively, the intermediate heat exchanger 16 may employ a counterflow heat exchanger in which the flow direction of the high-pressure refrigerant flowing through the high-pressure flow path 16 a is opposite to that of the intermediate-pressure refrigerant flowing through the intermediate-pressure flow path 16 b.

The outlet side of the intermediate-pressure refrigerant flow path 16 b of the intermediate heat exchanger 16 is coupled to the suction port side of the above high-pressure side compressor 11 (specifically, the high-pressure side compression mechanism 11 a) via a check valve (not shown). Thus, the high-pressure side compression mechanism 11 a of this embodiment sucks thereinto a mixture of the intermediate-pressure refrigerant flowing from the intermediate-pressure refrigerant flow path 16 b and the intermediate-pressure refrigerant discharged from the low-pressure side compressor 12.

In contrast, the outlet side of the high-pressure refrigerant flow path 16 a of the intermediate heat exchanger 16 is coupled to the inlet side of a low-pressure expansion valve 17. The low-pressure expansion valve 17 is a thermal expansion valve for decompressing and expanding the high-pressure refrigerant flowing from the radiator 13 into the low-pressure refrigerant. The low-pressure expansion valve 17 has the same basic structure as that of the intermediate-pressure expansion valve 15.

More specifically, the low-pressure expansion valve 17 has a temperature sensing portion disposed on the refrigerant outlet side of an evaporator 18 to be described later. The expansion valve 17 senses a degree of superheat of the refrigerant on the outlet side of the evaporator 18 based on the temperature and pressure of the refrigerant on the outlet side of the evaporator 18, and adjusts the opening degree of the valve (refrigerant flow rate) by a mechanical mechanism such that the superheat degree becomes a predetermined value preset.

The outlet side of the low-pressure expansion valve 17 is coupled to the refrigerant inflow side of the evaporator 18. The evaporator 18 is a heat exchanger for heat absorption which exchanges heat between the low-pressure refrigerant decompressed and expanded by the low-pressure expansion valve 17 and the blast air blown by a blower fan 18 a and circulating through the freezer to thereby exhibit the effect of heat absorption by evaporating the low-pressure refrigerant. The blower fan 18 a is an electric blower whose number of revolutions (amount of blast air) is controlled by a control voltage output from the refrigeration machine controller 20.

Further, the refrigerant outlet port of the evaporator 18 is coupled to the suction port side of the low-pressure side compressor 12 (specifically, the low-pressure side compression mechanism 12 a).

Next, an electric controller of this embodiment will be described later. The refrigeration machine controller 20 is comprised of the known microcomputer including a CPU for performing control processing or computation processing, and a storage circuit, such as a ROM or a RAM, for storing therein programs and data, an output circuit for outputting therefrom a control signal or a control voltage to a device for control, an input circuit into which a detection signal from each sensor is input, and a power supply circuit.

The output side of the refrigeration machine controller 20 is coupled to the above low-pressure side inverter 22, the high-pressure side inverter 21, the cooling fan 13 a, the blower fan 18 a, and the like as the devices to be controlled. The refrigeration machine controller 20 is adapted to control the operation of each of these devices to be controlled.

The refrigeration machine controller 20 includes a combination of respective control means for controlling the devices to be controlled. The respective structures (hardware and software) for controlling the operations of the devices to be controlled in the refrigeration machine controller 20 form the control means for controlling the respective devices for control.

In this embodiment, a first discharge capacity controller 20 a has the structure (hardware and software) for controlling the refrigerant discharge capacity of the low-pressure side compression mechanism 12 a by controlling the operation of the low-pressure side inverter 22. And, a second discharge capacity controller 20 b has the structure (hardware and software) for controlling the refrigerant discharge capacity of the high-pressure side compression mechanism 11 b by controlling the operation of the high-pressure side inverter 21.

Thus, the number of revolutions of the low-pressure side electric motor 12 b and the number of revolutions of the high-pressure side electric motor 11 b can be independently controlled by the first discharge capacity controller 20 a and the second discharge capacity controller 20 b. Obviously, the first and second discharge capacity controllers 20 a and 20 b may be composed of different controllers for the refrigeration machine controller 20.

In contrast, the input side of the refrigeration machine controller 20 is coupled to an outside air temperature sensor 23, an in-freezer temperature sensor 24, and the like. The outside air temperature sensor 23 serves as outside air temperature detection means for detecting an outside air temperature Tam of air outside the freezer (outdoor air) that exchanges heat with the high-pressure refrigerant at the radiator 13. The in-freezer temperature sensor 24 serves as in-freezer temperature detection means for detecting an air temperature Tfr of the blast air that exchanges heat with the low-pressure refrigerant at the evaporator 18. Detection signals from these sensors are input to the refrigeration machine controller 20.

The input side of the refrigeration machine controller 20 is coupled to an operation panel 30. The operation panel 30 is provided with an operation/stopping switch serving as request signal output means for outputting an operation request signal or a stopping request signal of the refrigeration machine, and a temperature setting switch serving as target temperature setting means for setting an in-freezer temperature (target cooling temperature) Tset. Operation signals from these switches are input to the refrigeration machine controller 20.

Next, the operation of the two-stage pressurizing refrigeration cycle device 10 with the above structure in this embodiment will be described below based on FIG. 2. First, FIG. 2 shows a flowchart of the control process performed by the refrigeration machine controller 20. The control process is started when the operation/stopping switch of an operation panel 30 is turned on (ON) to output an operation request signal.

First, in step S1, a flag, a timer, and the like are initialized. Then, in the following step S2, detection signals from the outside air temperature sensor 23 and the in-freezer temperature sensor 24, and an operation signal from a temperature setting switch or the like of the operation panel 30 are read in, and then an operation mode is determined according to a temperature Tset set by the temperature setting switch. Specifically, when the target cooling temperature Tset is −10° C. or more, a chilled mode is set in which fresh foods are refrigerated at a temperature for suppressing the degradation of freshness of the food. When the target cooling temperature Tset is lower than −10° C., a frozen mode is set in which freezing is performed.

Subsequently, the operation proceeds to step S3, in which a control mode is determined. The control mode is common to the chilled mode and the frozen mode, and thus a description of the control mode for each operation mode will be omitted below.

Specifically, in step S3, when a difference in temperature ΔT obtained by subtracting a target cooling temperature Tset set by the temperature setting switch from the air temperature Tfr read in step S2 is larger than a predetermined reference temperature difference ΔKT, the large capacity is determined to be necessary. In contrast, when the temperature difference ΔT is equal to or less than the predetermined reference temperature difference ΔKT, it is determined that the temperature of the inside of the freezer approaches the preset temperature Tset, and thus the detailed control of the capacity is necessary.

In most cases, immediately after start-up of the refrigeration machine, the temperature of the inside of the freezer as the space for cooling is higher than the target cooling temperature Tset. Thus, the temperature difference ΔT used in this embodiment is a value obtained by subtracting the target cooling temperature Tset from the air temperature Tfr. However, an absolute value obtained by subtracting the air temperature Tfr from the target cooling temperature Tset may be used as the temperature difference ΔT.

When the large capacity is determined to be necessary in step S3, the operation proceeds to step S4 in which the operation in a cool-down mode is performed. In step S4, the numbers of revolutions of the high-pressure side electric motor 11 b and the low-pressure side electric motor 12 b are determined such that the refrigerant discharge capacity of the low-pressure side compressor 12 and the refrigerant discharge capacity of the high-pressure side compressor 11 are substantially maximized.

In the following step S5, the control states of other devices to be controlled in the cool-down mode of the refrigeration machine are determined. For example, the numbers of revolutions of the cooling fan 13 a and the blower fan 18 a are determined such that the blowing capacities thereof are substantially maximized. Then, the operation proceeds to step S9.

In contrast, when the detailed control of the capacity of the refrigeration machine is determined to be necessary in step S3, the operation proceeds to step S6, in which the operation in a capacity control mode is performed. In step S6, a refrigerant discharge capacity of the low-pressure side compressor 12 is determined based on the detection signal and the operation signal read in the present step S2.

More specifically, the number of revolutions of the low-pressure side electric motor 12 b, that is, the number of revolutions N2 of the low-pressure side compression mechanism 12 a is determined based on elements, including a deviation, an integral, and a differentiation of the control temperature and the preset temperature in step S6.

In the following step S7, a refrigerant discharge capacity of the high-pressure side compressor 11 is determined based on the refrigerant discharge capacity of the low-pressure side compressor 12 determined in step S6.

Specifically, in step S7, the number of revolutions N1 of the high-pressure side compression mechanism 11 a is determined such that an effective capacity ratio determined by the following formula F1 is within a predetermined reference range represented by the following formula F2:

Effective Capacity Ratio=N2×V2/N1×V1  (F1)

1≦N2×V2/N1×V1≦3  (F2)

in which V1 is a discharge capacity of the high-pressure side compression mechanism 11 a, N1 is the number of revolutions of the high-pressure side compression mechanism 11 a, V2 is a discharge capacity of the low-pressure side compression mechanism 12 a, and N2 is the number of revolutions of the low-pressure side compression mechanism 12 a.

In the following step S8, the control states of other devices to be controlled are determined. For example, the number of revolutions of each of the cooling fan 13 a and the blower fan 18 a is determined such that its blowing capacity is increased with increasing number of revolutions N2 of the low-pressure side compression mechanism 12 a determined in step S6, and then the operation proceeds to step S9.

Then, in step S9, a control signal is output from the refrigeration machine controller 20 to the device to be controlled, which is coupled to the output of the controller, so as to obtain the control state determined in steps S4 to S8, and then the operation proceeds to step S10.

In step S10, when a stopping request signal is output from the operation panel 30 to the refrigeration machine controller 20, the operation of each device to be controlled is stopped to thereby completely stop the entire system of the refrigeration machine. In contrast, when the stopping request signal is not output, the operation returns to S2 after the expiration of a predetermined control cycle τ.

Thus, when the operation/stopping switch of the operation panel 30 is turned on, the high-pressure side compressor 11 in the two-stage pressurizing refrigeration cycle device 10 sucks, compresses, and discharges a mixture of an intermediate-pressure refrigerant discharged from the low-pressure side compressor 12 and an intermediate-pressure refrigerant flowing from the intermediate-pressure refrigerant flow path 16 b of the intermediate heat exchanger 16.

And, the high-temperature and high-pressure refrigerant discharged from the high-pressure side compressor 11 flows into the radiator 13, and exchanges heat with the air outside the freezer blown by the cooling fan 13 a to be thereby cooled. The flow of the high-pressure refrigerant flowing from the radiator 13 is branched by the branching portion 14. Then, the high-pressure refrigerant flowing from the branching portion 14 into the intermediate-pressure expansion valve 15 is decompressed and expanded until it is converted into the intermediate-pressure refrigerant.

At this time, the throttle opening of the intermediate-pressure expansion valve 15 is adjusted such that the degree of superheat of the refrigerant at the outlet of the intermediate-pressure refrigerant flow path 16 b of the intermediate heat exchanger 16 becomes a predetermined value preset. Further, the intermediate-pressure refrigerant decompressed by the intermediate-pressure expansion valve 15 flows into the intermediate-pressure refrigerant flow path 16 b of the intermediate heat exchanger 16. The refrigerant exchanges heat with the high-pressure refrigerant flowing from the branching portion 14 into the high-pressure refrigerant flow path 16 a of the intermediate heat exchanger 16 to be thereby heated, and then is sucked into the high-pressure side compressor 11.

In contrast, the high-pressure refrigerant flowing from the branching portion 14 into the high-pressure refrigerant flow path 16 a of the intermediate heat exchanger 16 is cooled by the intermediate heat exchanger 16. The high-pressure refrigerant flowing from the high-pressure refrigerant flow path 16 a flows into the low-pressure expansion valve 17, and decompressed and expanded until it becomes the low-pressure refrigerant. At this time, the throttle opening of the low-pressure expansion valve 17 is adjusted such that the degree of superheat of the refrigerant at the outlet of the evaporator 18 becomes the predetermined value preset.

Further, the low-pressure refrigerant decompressed by the low-pressure expansion valve 17 flows into the evaporator 18, and absorbs heat from the blast air circulating therethrough and blown by the blower fan 18 a to evaporate itself. Thus, the blast air blown into the freezer as a space for cooling is cooled. The refrigerant flowing from the evaporator 18 is sucked into the low-pressure side compressor 12.

The two-stage pressurizing refrigeration cycle device 10 of this embodiment is operated as mentioned above to thereby form the above economizer refrigeration cycle device, so that it can not only improve the compression efficiency of the high-pressure side compression mechanism, but also can exhibit the following excellent effects.

First, in this embodiment, the refrigerant discharge capacity of the low-pressure side compression mechanism 12 a is determined based on the outside air temperature Tam, the air temperature Tfr, and the preset temperature Tset, and further the refrigerant discharge capacity of the high-pressure side compression mechanism 11 a is determined based on the determined refrigerant discharge capacity of the low-pressure side compression mechanism 12 a. Thus, the refrigerant discharge capacities of the respective compression mechanisms 11 b and 12 b can be easily determined.

At this time, the refrigerant discharge capacity of the high-pressure side compression mechanism 11 a is determined such that the effective capacity ratio satisfies the above formula F2. Thus, the coefficient of performance (COP) of the cycle can be improved with the simple structure without the need for the pressure detection means for detecting the high-pressure side refrigerant pressure, the intermediate-pressure side refrigerant pressure, or the low-pressure side refrigerant pressure, and under excessively easy control.

This will be described below in detail using FIG. 3. FIG. 3 shows a graph of a change in COP ratio relative to a change in effective capacity ratio. FIG. 3(a) is a graph on the condition A of an outside air temperature Tam=38° C., and a preset temperature Tset=−10° C. FIG. 3( b) is a graph on the condition B of an outside air temperature Tam=10° C., and a preset temperature Tset=−25° C.

The term “COP ratio” as used herein means the ratio of the COP obtained when the intermediate refrigerant pressure is set to a predetermined value different from a geometric mean between the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure to the COP of the two-stage pressurizing refrigeration cycle device 10 of this embodiment.

As can be seen from FIG. 3, the COP ratio has a peak at an effective capacity ratio in a range of not less than 1 nor more than 3 on any condition. This means that by setting the effective capacity ratio in a range of not less than 1 nor more than 3, the intermediate refrigerant pressure can approach the geometric mean between the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure.

Thus, the two-stage pressurizing refrigeration cycle device 10 of this embodiment can improve the COP with the simple structure under the very easy control. Even on any condition, the peak of the COP ratio exists at an effective capacity ratio close to 2. In control step S7, the COP can be further improved by setting the effective capacity ratio in a range of not less than 1.5 nor more than 2.5.

Like this embodiment, in the refrigeration cycle applied to the refrigeration machine, for example, a difference in pressure between the high-pressure side refrigerant pressure and the low-pressure side refrigerant pressure becomes large as compared to the refrigeration cycle applied to an air conditioner, thereby tending to increase the consumption power of the compressor. Thus, the improvement of the COP is very useful in the refrigeration cycle applied to the refrigeration machine.

In this embodiment, regardless of the refrigerant discharge capacities of the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a, the throttle opening of the intermediate-pressure expansion valve 15 is adjusted such that the refrigerant at the outlet of the intermediate-pressure refrigerant flow path 16 b of the intermediate heat exchanger 16 has the adequate superheat degree. This arrangement can avoid the problem of liquid compression of the high-pressure side compression mechanism 11 a. Further, the superheat degree of the low-pressure expansion valve 17 is adjusted such that the refrigerant at the outlet of the evaporator 18 has the adequate superheat degree, and thereby it can also avoid the problem of liquid compression of the low-pressure side compression mechanism 12 a.

Thus, even the simple structure can improve the reliability of the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a, that is, the reliability of the entire two-stage pressurizing refrigeration cycle device.

In this embodiment, the refrigerant discharge capacity of the low-pressure side compression mechanism 12 a is determined based on the outside air temperature Tam or the like, so that the refrigerant evaporation pressure of the evaporator 18 can be directly determined based on the outside air temperature Tam or the like. Thus, the air temperature Tfr of the blast air blown into the freezer can easily approach the preset temperature Tset.

The two-stage pressurizing refrigeration cycle device 10 of this embodiment includes the intermediate heat exchanger 16, whereby the intermediate-pressure refrigerant flowing from the intermediate-pressure expansion valve 15 can be heated by the high-pressure refrigerant branched by the branching portion 14 to be easily converted into the gas-phase refrigerant. As a result, the reliability of the two-stage pressurizing refrigeration cycle device can be more surely improved.

The high-pressure refrigerant branched by the branching portion 14 can be cooled by the intermediate-pressure refrigerant flowing from the intermediate-pressure expansion valve 15, thereby enlarging a difference in enthalpy between the refrigerant at an inlet of the evaporator 18 and the refrigerant at an outlet thereof to thereby increase the refrigeration capacity exhibited by the evaporator 18. As a result, the COP of the two-stage pressurizing refrigeration cycle device can be further improved.

In this embodiment, when the refrigeration cycle device is determined not to be in the state immediately after the start-up of the refrigeration machine in control step S3, the refrigerant discharge capacity of the high-pressure side compression mechanism 11 a is determined based on the refrigerant discharge capacity of the low-pressure side compression mechanism 12 a. Thus, immediately after the start-up of the refrigeration machine, the refrigerant discharge capacity of the low-pressure side compression mechanism 12 a and the refrigerant discharge capacity of the high-pressure side compression mechanism 11 a are substantially maximized, so that the operation mode for rapidly cooling the space for cooling can be performed.

In this embodiment, the fixed displacement compression mechanism is employed as the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a, so that the discharge capacities V1 and V2 of the compression mechanisms can be constant. Thus, after determining the number of revolutions N2 of the low-pressure side compression mechanism 12 a, the effective capacity ratio can be easily in a desired range only by adjusting the number of revolutions N1 of the high-pressure side compression mechanism 11 a.

Second Embodiment

As shown in the entire configuration diagram of FIG. 4, a second embodiment will be described below by way of example, which includes the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a formed of a variable displacement compression mechanism, unlike the first embodiment. Further, in this embodiment, the high-pressure side electric motor 11 b and the low-pressure side electric motor 12 b are removed, so that both compression mechanisms 11 a and 12 a are rotatably driven by a common electric motor 19. In FIG. 4, the same or equivalent parts as those of the first embodiment are represented by the same reference numeral.

More specifically, in this embodiment, the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a employ a swash plate type variable displacement compression mechanism. The swash plate type variable displacement compression mechanism changes a control pressure Pc of a swash plate chamber by a swash plate compressor to thereby vary an inclined angle of a swash plate so as to change a stroke of a piston, whereby the discharge capacity is continuously changed in a range of substantially 0 to 100%.

The control pressure Pc of the inside of a corresponding swash plate chamber in each of the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a is adjusted by changing an opening degree of each of electromagnetic capacity control valves 11 c and 12 c to change a rate of introduction of each of the high-pressure refrigerant and the low-pressure refrigerant to be introduced into the swash plate chamber. The electromagnetic capacity control valves 11 c and 12 c have the operation thereof controlled by the control current output from first and second discharge capacity controllers 20 a and 20 b of the refrigeration machine controller 20.

An electric motor 19 is an AC motor whose operation (number of revolutions) is controlled by alternating current output from an inverter 25, like the high-pressure side electric motor 11 b and the low-pressure side electric motor 12 b. An inverter 25 outputs the alternating current with a frequency corresponding to a control signal output from the refrigeration machine controller 20.

Further, the rotation driving force output from the electric motor 19 of this embodiment is transferred to both compression mechanisms 11 a and 12 a via a pully and a belt. Thus, a ratio of the number of revolutions N2 of the low-pressure side compression mechanism 12 a to the number of revolutions N1 of the high-pressure side compression mechanism 11 a, that is, N2/N1 in this embodiment is set to a constant value. In this embodiment, the ratio of the number of revolutions N2/N1 is set to substantially 1, so that the number of revolutions N2 of the low-pressure side compression mechanism 12 a is substantially equal to the number of revolutions N1 of the high-pressure side compression mechanism 11 a.

Other structures and operations of this embodiment are the same as those of the first embodiment. Thus, like the first embodiment, the two-stage pressurizing refrigeration cycle device 10 of this embodiment can also improve the coefficient of performance (COP) of the cycle with a simple structure under excessively easy control. Further, the above simple structure of this embodiment can improve the reliability of the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a, that is, the reliability of the entire two-stage pressurizing refrigeration cycle device.

In this embodiment, a variable displacement compression mechanism is employed as the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a, whereby the discharge capacities V1 and V2 of the compression mechanisms 11 a and 12 a can be independently changed. Even when the numbers of revolutions N1 and N2 of both the compression mechanisms 11 a and 12 a are the same value, the effective capacity ratio (N2×V2/N1×V1) can be easily changed to a desired value.

Both the compression mechanisms 11 a and 12 a can be driven by a common driving source (electric motor 19), thereby making the cycle structure much simpler.

Other Embodiments

The present invention is not limited to the above embodiments, and various modifications and changes can be made to those embodiments without departing from the spirit and scope of the invention.

(1) Although in the above embodiments, the cycle structure of the two-stage pressurizing refrigeration cycle device employs the intermediate heat exchanger 16, the invention is not limited thereto. For example, the intermediate heat exchanger 16 may be removed, and an intermediate gas-liquid separator for separating the refrigerant flowing from the intermediate pressure expansion valve 15, into liquid and gas phases may be provided.

The gas-phase refrigerant separated by the intermediate gas-liquid separator may be sucked into the high-pressure side compressor 11. In this case, the intermediate-pressure expansion valve 15 may be removed, and instead, a fixed throttle may be employed. Further, the branching portion 14 may be removed to allow the liquid-phase refrigerant separated by the intermediate gas-liquid separator to flow into the low-pressure expansion valve 17, so as to construct an economizer refrigeration cycle device.

(2) In the above embodiment, in control step S6 shown in FIG. 2, a refrigerant discharge capacity of the low-pressure side compressor 12 is determined based on the outside air temperature Tam or the like. In control step S7, a refrigerant discharge capacity of the high-pressure side compressor 11 is determined based on the refrigerant discharge capacity of the low-pressure side compressor 12, by way of example. Alternatively, likewise, a refrigerant discharge capacity of the high-pressure side compressor 11 may be determined in control step S6, and a refrigerant discharge capacity of the low-pressure side compressor 12 may be determined in control step S7.

In control step S6, the refrigerant discharge capacity of the low-pressure side compressor 12 is determined based on the outside air temperature Tam, the air temperature Tfr, and the preset temperature Tset, by way of example. Alternatively, the refrigerant discharge capacity of the low-pressure side compressor 12 may be determined using at least one of the outside air temperature Tam, the air temperature Tfr, and the preset temperature Tset.

(3) In the above embodiments, the thermal expansion valve is employed as the intermediate-pressure expansion valve 15 and the low-pressure expansion valve 17, by way of example. Alternatively, an electric expansion valve may be used as the intermediate-pressure expansion valve 15 and the low-pressure expansion valve 17.

For example, in addition, detection means may be provided for detecting the temperature and pressure of the refrigerant on the outlet side of the intermediate-pressure refrigerant flow path 16 b, and the operation of the intermediate-pressure expansion valve 15 may be controlled such that the degree of superheat of the refrigerant on the outlet side of the intermediate-pressure refrigerant flow path 16 b becomes a predetermined value preset. Further, another detection means may be added for detecting the temperature and pressure of the refrigerant on the outlet side of the evaporator 18, and the operation of the low-pressure side expansion valve 17 may be controlled such that the superheat degree of the refrigerant on the outlet side of the evaporator 18 becomes a predetermined value preset.

(4) In the above embodiments, the two-stage pressurizing refrigeration cycle device 10 of this embodiment is applied to the refrigeration machine by way of example, but the invention is not limited thereto. For example, the refrigeration cycle device may be applied to an air conditioner, a refrigerator, and the like. Alternatively, the refrigeration cycle device may be applied to a refrigerating and freezing container of a mobile object (vehicle, ship) and the like.

Stationary air conditioner, refrigerator, and freezer are easy to obtain energy for driving of the compressors 11 and 12 from a commercial power source or the like. However, this kind of refrigerating and freezing container applied to the mobile object has a limited driving energy. Thus, the improvement of the COP achieved by the two-stage pressurizing refrigeration cycle device 10 in the invention is very useful.

(5) In the above control step S3, it is determined whether the large capacity is necessary or not, by comparing the difference in temperature ΔT with the reference temperature difference ΔKT. The determination way is not limited thereto.

For example, after the operation/stopping switch is turned on (at the time of ON), first, the numbers of revolutions of both compression mechanisms 11 a and 12 a are determined such that the difference between the air temperature Tfr and the target cooling temperature Tset is reduced. In such control, when the amount of change ΔTfr in air temperature Tfr per unit time is more than a predetermined reference temperature change amount ΔKTfr, the refrigeration machine is determined to be in the state immediately after the start-up of the refrigeration machine. When the ΔTfr is equal to or less than the predetermined reference temperature change amount ΔKTfr, the refrigeration machine may be determined to be in a normal operation state.

(6) In use of an azeotropic refrigerant or dummy azeotropic refrigerant as the refrigerant, a difference in temperature between the refrigerant at the inlet of the intermediate-pressure refrigerant flow path 16 b and the refrigerant at the outlet of the intermediate-pressure refrigerant flow path 16 b may be detected. Then, the opening degree of the intermediate-pressure expansion valve 15 (flow rate of the refrigerant) may be adjusted such that the temperature difference becomes a predetermined value preset.

As a refrigerant temperature may be used a surface temperature of a refrigerant pipe connecting the intermediate-pressure refrigerant flow path 16 b to another component.

(7) In the above first embodiment, in steps S5 and S7, other devices except for the low-pressure side compressor and the high-pressure side compressor (cooling fan 13 a and blower fan 18 a) are controlled according to the control mode, but may be controlled according to the operating mode. For example, the devices may be controlled such that the blowing capacities of the cooling fan 13 a and the blower fan 18 a are substantially maximized in the chilled mode, and such that the blowing capacities of the cooling fan 13 a and the blower fan 18 a become low in the frozen mode. (8) In the above embodiment, in step S6, the number of revolutions of the low-pressure side electric motor 12 b is controlled using the control temperature and the preset temperature, that is, under the so-called PID control. Alternatively, the number of revolutions of the low-pressure side electric motor 12 b, that is, the number of revolutions N2 of the low-pressure side compression mechanism 12 a may be determined based on the outside air temperature Tam, the air temperature Tfr, and the preset temperature Tset with reference to a control map pre-stored in a storage circuit of the refrigeration machine controller 20 in the following way. The number of revolutions N2 may be determined such that the refrigerant discharge capacity of the low-pressure side compressor 12 is increased with increasing outside air temperature Tam, increasing air temperature Tfr, or decreasing preset temperature Tset. (9) In the above second embodiment, both the high-pressure side compression mechanism 11 a and the low-pressure side compression mechanism 12 a are driven using one electric motor 19 as driving means by way of example. Alternatively, different driving means may be used for the respective compression mechanisms 11 a and 12 a, and an engine (internal combustion engine) may be used as the driving means. 

1-8. (canceled)
 9. A two-stage pressurizing refrigeration cycle device comprising: a low-pressure side compression mechanism, which compresses a low-pressure refrigerant into a first intermediate-pressure refrigerant, and discharges the first intermediate-pressure refrigerant therefrom; a high-pressure side compression mechanism, which compresses the first intermediate-pressure refrigerant discharged from the low-pressure side compression mechanism into a high-pressure refrigerant to discharge the high-pressure refrigerant therefrom; a radiator, which exchanges heat between outdoor air and the high-pressure refrigerant discharged from the high-pressure side compression mechanism, to dissipate heat from the refrigerant; an intermediate-pressure expansion valve, which decompresses and expands the high-pressure refrigerant flowing from the radiator into a second intermediate-pressure refrigerant to flow the second intermediate-pressure refrigerant into a suction side of the high-pressure side compression mechanism; a low-pressure expansion valve, which decompresses and expands the high-pressure refrigerant flowing from the radiator into a low-pressure refrigerant; an evaporator, which evaporates the low-pressure refrigerant decompressed and expanded by the low-pressure expansion valve by exchanging heat with air to be blown into a space for cooling, and is connected to a suction side of the low-pressure side compression mechanism to cause the evaporated refrigerant to flow into the suction side of the low-pressure side compression mechanism; a first discharge capacity controller configured to determine a refrigerant discharge capacity of at least the low-pressure side compression mechanism, such that the refrigerant discharge capacity is increased in accordance with an increase in at least one of an outside air temperature of the outdoor air for exchanging heat with the high-pressure refrigerant at the radiator and an air temperature of air for exchanging heat with the low-pressure refrigerant at the evaporator; and a second discharge capacity controller configured to determine a refrigerant discharge capacity of the high-pressure side compression mechanism based on the determined refrigerant discharge capacity of the-low-pressure side compression mechanism, wherein the second discharge capacity controller determines the refrigerant discharge capacity of the high-pressure side compression mechanism, such that an effective capacity ratio defined as N2×V2/N1×V1 is within a predetermined reference range, when V1 is a discharge capacity of the high-pressure side compression mechanism, N1 is the number of revolutions of the high-pressure side compression mechanism, V2 is a discharge capacity of the low-pressure side compression mechanism, and N2 is the number of revolutions of the low-pressure side compression mechanism.
 10. The two-stage pressurizing refrigeration cycle device according to claim 9, wherein the intermediate-pressure expansion valve decompresses and expands one high-pressure refrigerant flow branched at a branching portion in which the high-pressure refrigerant flowing from the radiator is branched, and the low-pressure expansion valve decompresses and expands an other high-pressure refrigerant flow branched at the branching portion, said refrigeration cycle device further comprising an intermediate heat exchanger in which heat is exchanged between the low-pressure refrigerant decompressed and expanded by the intermediate-pressure expansion valve and the other high-pressure refrigerant branched by the branching portion.
 11. The two-stage pressurizing refrigeration cycle device according to claim 9, wherein when an absolute value of a difference in temperature between the air temperature at the evaporator and a target cooling temperature of the space for cooling is equal to or less than a predetermined reference temperature difference, the second discharge capacity controller determines the refrigerant discharge capacity of the high-pressure side compression mechanism based on the refrigerant discharge capacity of the low-pressure side compression mechanism.
 12. The two-stage pressurizing refrigeration cycle device according to claim 11, wherein when the air temperature at the evaporator is higher than the target cooling temperature, and when the absolute value of the difference in temperature between the air temperature at the evaporator and the target cooling temperature of the space for cooling is equal to or less than the predetermined reference temperature difference, the second discharge capacity controller determines the refrigerant discharge capacity of the high-pressure side compression mechanism based on the refrigerant discharge capacity of the low-pressure side compression mechanism.
 13. The two-stage pressurizing refrigeration cycle device according to claim 9, wherein each of the high-pressure side compression mechanism and the low-pressure side compression mechanism is configured by a fixed displacement compression mechanism having a fixed discharge capacity, said refrigeration cycle device further comprising: a high-pressure side electric motor, which rotatably drives the high-pressure side compression mechanism; and a low-pressure side electric motor, which rotatably drives the low-pressure side compression mechanism, wherein the number of revolutions of the high-pressure side electric motor and the number of revolutions of the low-pressure side electric motor are independently controllable.
 14. The two-stage pressurizing refrigeration cycle device according to claim 9, wherein each of the high-pressure side compression mechanism and the low-pressure side compression mechanism is configured by a variable displacement compression mechanism having a variable discharge capacity, and the discharge capacity of the high-pressure side compression mechanism and the discharge capacity of the low-pressure side compression mechanism are independently controllable.
 15. The two-stage pressurizing refrigeration cycle device according to claim 9, wherein the second discharge capacity controller determines the refrigerant discharge capacity of the high-pressure side compression mechanism, such that the effective capacity ratio satisfies the following formula: 1≦N2×V2/N1×V1≦3. 